Vibratory drill apparatus



Inventors Nick I). Diamantides [561 2517 14th St., Cuyahoga Falls,44223; William L. Rinks, 2449 Kensington Ave., 1 1 12 498 Bath, Ohio44210 l365l Appl. No. 850,682 29463l4 Filed May 15, 1969 295335lPatented Oct. 6, 1970 3,307,641 3,371,726

References Cited UNITED STATES PATENTS 10/1914 Vanes 4/1938Heaston........... 7/1960 Nast 9/1960 Bodine 3/1967 Wiggins 3/1968Bouyoucosu Primary Examiner-Nile C. Byers, .lr. AttorneyN. D.Diamantides VIBRATORY DRILL APPARATUS l75/56X l75/56X l75/56X 175/56175/56 l75/56X ABSTRACT: The subject matter of this invention is a rock58 claimsza Drawing Figs drill apparatus whose working member is driven,both, to a 0.8. CI..... 175/56 high frequency longitudinal vibration andto indexing through Int. Cl. E2lb 7/04 rotation, and which is powered bythe pressurized fluid Field of Search 175/56, customarily used inremoving the rock debris from the drilled 107; 253/22, 33 hole.

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s Q ,0 I f nu i w 11 I40. I2a L 40 E- 12b /l4b l3b I I O l Patented Oct.6, W70

Sheet INVENTORS N. D. DIAMANTIDES W.L. HINKS Patented Oct. 6, 19703,532,114

Sheet 2 of 9 N. D. DIAMANTIDES k BY w.1 HJNKS 345 Patented Oct. 6, 1970Sheet INVENTORS N. D. DIA MANTIDES BY W. L. HINKS Patented 0a.. 6, 19703,532,174

Sheet {5 of9 33E INVENTORS 515 Be. v N. 0. DIAMANTIDES BY W.L. HINKSPatented 0a,. .6, 1970 3,532,174

- Sheet 6 019 V 5,, gas ii INVENT 22 N. D. D/AMANT S BY W. L'. HINKSPatented Oct. 6, 1970 Sheet 7 INVENTORS N. D, DIAMANTIDES W.L. H/NKSPatented Oct. 6, 1970 Sheet of 9 W. l- HINKS VIBRATORY DRILL APPARATUSThis application is a continuation of Ser. No. 734,048 filed June 3,1968, now abandoned.

This invention refers in general to the field of rock penetrating tools,and in particular to a rock drill.

It is well known to those versed in this art that rock penetration forthe purposes of ground sampling, mining, quarrying, oil exploitation,and tunnel construction faces many technical and economic problemsdependent both on the geological forplicity of handling and repairing.

A further object of the invention is to provide a rock drill requiring aminimum of peripheral and auxiliary equipment and a limited on-the-siterig structure.

Still another object is to provide a rock drill whose function .rests ona rotary-vibratory movement of its working head, and

is powered by a compressed fluid source.

A further object is a rock drill in which vibratory movement is impartedto the working head of the drill by a mechanism involving no solid partssliding on one another, but, instead, allowing a minimum number of suchparts to move with respect to each other by means of a specialdeformable connecting member.

Additional objects of the invention will become apparent upon reading ofthe following specification, considered and interpreted in the light ofthe accompanying drawings in which:

FIG. I is a schematic drawing showing the four principal components oftheir rock drill and their preferred mechanical interconnection.

FIG. 2 is a cross-sectional view of one configuration of the rock drillequipped with a tricone type bit and showing the arrangement of laminatebearings connecting the principal drill components.

FIG. 2a is a partial cross-sectional view of the vibrator and headshowing the use of Bellville type springs in place of laminate bearings.

FIG. 2b is a partial cross section of the vibrator and head showing theuse of a resilient elastomer ring in place of laminate bearings.

FIG. 3 is a cross-sectional view of the vibrator and head, equipped witha fish tail or drag type bit, illustrating fluidic oscillator outlets intheir association with the annular piston chambers for the generation ofthe fluid forces that set the vibrator and head in reciprocating motion.

FIG. 4 is a cross-sectional view of the vibrator and head illustratingthe fluidic oscillator arrangement and its feedback paths emanating fromthe piston chambers.

FIG. 4a is a cross-sectional view of the head and vibrator illustratingthe use of individual laminate bearings at each piston chamber.

FIG. 5 is a schematic diagram of the fluidic oscillator aiding theexplanation of the basic principles of its operation.

FIG. 6 is a cross-sectional view of the vibrator and head showing amechanical oscillator arrangement and its association with the annularpiston chambers for the generation of the fluid forces that set thevibrator and head in reciprocating motion.

FIG. 6a is a detailed drawing of an alternate arrangement of the jetnozzles and passages of the mechanical oscillator and vibrator.

FIG. 6b is a detailed drawing of still another arrangement of jetnozzles and passages of the mechanical oscillator and vibrator.

FIG. 60 is a detailed drawing showing an alternate arrangement of spentfluid paths through the mechanical oscillator, vibrator and head.

FIG. 6d is a detailed drawing showing the connection of the mechanicaloscillator to the vibrator and head by means of laminate bearings and aviscous damper respectively.

FIG. 6e is a detailed drawing showing still another connection of themechanical oscillator to the vibrator and head with the positions of thelaminate bearings and viscous damper of the previous figure reversed.

FIG. 7 is a cross-sectional view of the vibrator showing theincorporation of a fluidic oscillator for imparting a reciprocatingmotion to the mechanical oscillator.

FIG. 8 is a cross-sectional view of the vibrator and head showing anarrangement of the fluidic oscillator conducive to introducing amotional feedback to the function of the oscilla- I01.

FIG. 9 is a cross-sectional view of the vibrator and head showing anarrangement whereby reaction jets are used to generate the vibratoryforces required by the present drill.

FIG. 10 is a cross-sectional view of the rock drill showing a mechanicalinterconnection between head, vibrator, and drill stem different fromthat of FIG. 1.

FIG. 11 is a cross-section view of the rock drill showing an arrangementof its parts whereby the bit is attached to the vibrator and the sternadapter to the head.

FIG. 12 is a partially cross-sectional elevated view of the vibrator andhead illustrating the incorporation of a hydraulic wheel for thegeneration of the torque that rotates the head.

FIG. 12a is a plane cross-sectional view of the vibrator and head alongthe line Ila-12a of FIG. 12.

FIG. 13 is a partially cross-sectional elevated view of the vibrator andhead illustrating a special laminated bearing arrangement forimplementing a stepped rotation of the head and bit as well as axialvibration of it.

FIG. 14 is a cross-sectional view of the rock drill showing stillanother mechanical interconnection between head, vibrator and drillstem.

FIG. 15 is a schematic diagram of the equivalent mechanical circuit ofthe arrangement of the :rock drill components shown in FIGS. 2, 3, 4, 8,and 9.

FIG. 16 is a schematic diagram of the equivalent mechanical circuit ofthe arrangement of the rock drill components shown in FIG. 10.

FIG. 17 is a schematic diagram of the equivalent mechanical circuit ofthe arrangement of the :rock drill components shown in FIGS. 6 and 7. 7

FIG. 18 is a schematic diagram of the equivalent mechanical circuit ofthe arrangement of the :rock drill components shown in FIG. 14.

FIG. 19 is a typical force-displacement characteristic associated withthe crushing of a rock by drill.

Referring now to the preferred mechanical configuration of FIG. 1, itwill be seen that the rock drill of the present invention comprises incombination the stern adapter 10, which is the lower end part of thedrill stem 10a conventionally used in rock drilling; the vibrator 20;the head 40; and the bit 50. Details of their mechanical interconnectionmay be seen in FIGS. 3 and 6, in which two of the drills configurationsare shown and in FIG. 2 which applies to both. In all three cases themassive vibrator 20 is attached resiliently in the axial direction tothe stem adapter 10 by means of the interposed laminate elastomericbearings or bushings 11. A telescoping arrangement, with annularbearings is preferred, as shown. The stem adapter 10 is indicatedexterior to the top of vibrator 20, but obviously the top of vibrator 20could be made the exterior part and the stem adapter 10 interior, withthe annular bearings 11 interposed between. Two axially separatedbearing assemblies are shown, but there could be instead one longbearing assembly. The total amount of rubber in shear must be able tosustain the axial load applied by the stem. Similarly, the massive head40 is attached in an axially resiliently stiff manner to the vibrator 20preferably by means of the laminate elastomeric bearings 12a and 12b.One or all bearings 11, 12a, 12b may be of the rubber sleeve, instead oflaminate type. In both configurations the sections 13 and] 13b of thevibrator 20 have a polygonal cross section (in a plane perpendicular tothe longitudinal axis) as do the respectively opposed sections 14 and14b of the head 40 and stem adapter 10. By contrast, the laminatebearing 12a may be made conical and split in two halves radially tofacilitate assembling the vibrator-head system. The polygonal crosssections, in fact any noncircular shape, enable torque to be transmittedfrom the stem adapter 10 to the vibrator 20 and from the latter to thehead 40 effectively without windup. A round cross section except for oneor more flat sides could accomplish the same purpose. The bearings 11,12a, and 12b have the additional function of serving as seals betweenthe various sections of the flow path of the pressurized drill fluid.Finally, the bit 50 is affixed rigidly or in a drivable manner to thehead 40 by means of the tapered threaded sections 51 (bit not shown inFIG. 6). The bit is shown to be of the tricone type in FIG. 2 and of thefish tail or drag type in FIG. 3. Obviously, a coupler could beinterposed between head and bit for the purpose of increasing mass orlength or inserting compliance or damping between head and bit. in thesame two FIGS. 2 and 3 a shoulder 5 is shown interior to central channel7 of the stem adapter 10 that serves as a mechanical stop againstaccidental overloading of the bearings 11.

In the aforementioned designs shown in FIGS. 2. 3, 4, and 6, the massivevibrator is interior to the massive head 40. However, as disclosed laterthese relative positions may be interchanged. Since either configurationmay be used, it is deemed appropriate to use the terms inner massivemember and outer massive member when referring to both configurationscollectively.

The laminate elastomeric bearings constitute an important member of thepreferred configuration on account of their proven unique mechanicalbehavior. A bearing of this kind consists of a stack of thin metallaminates interleaved and adhered together by thin alternating layers ofelastic rubber or other rubberlike material. Such a layer of rubberlikematerial bonded between metal laminae can withstand high compressiveloads applied by the metal layers, it being sufficiently thin as to berestrained from substantially flowing sidewise by its adhesion to themetal. Elastomeric bearings, however, are capable of a deformation inshear parallel to the laminae accompanied by shear stress proportionalto the deformation. Thus the elastomeric bearing behaves in a softspringlike fashion in a plane parallel to the laminae, and its springrate may be adjusted as desired through proper selection of thedimensions of the stack and the characteristics of the elastomer. Thisspring rate, coplanar with the laminae, is little effected by anycompressive forces that may be applied perpendicularly to the layers.

The foregoing mode of bearing operation is the one under which thebearings 11 between stem adapter 10 and vibrator 20 are indicated tofunction with layers parallel to the longitudinal axis. However, this isnot necessarily the case with the bearings 12a and 12b between vibrator20 and head 40, for the latter bearings have to be much stiffer axiallythan the bearings 11 and yet able to withstand the severe fatiguestresses necessitated by the present drill design. To accomplish thiswithin small dimensions and without the need for a very stiff elastomer,the laminate bearings 12a, 12b are set with their load faces at an anglewith respect to the motion 2: this produces not only shear stressesparallel to the laminae but, also, compression perpendicular to thelaminae, with the net result that the equivalent spring constant isincreased, by orders of magnitude if necessary, as needed for a highfrequency oscillating system. The sloped bearings 12a and 1212 willgenerally be preloaded against one another so that, within theoscillation amplitude experienced, neither may become completelydecompressed. An opposingly preloaded pair of bearings is thus providedplacing the vibrator in compression. The vibrator will be subjected totension instead if the sloping angles of both bearings is reversed.

It should be emphasized that by their nature, elastomeric bearings arecapable of prolonged life with proper design,

without lubrication and without fatigue and failure, withstanding severevibratory deformations under considerable static loading. Regarding thetwo constituents of the laminate elastomeric beaiings itshotlldbelipderstood that the term metafas used herein encompasses any ofi'f'ir'iBM of metals, or eyen a nonmetallic material, that is characterized bya high compression resistence, as well as high tensile strength;similarly, when the term elastomeric or rubberlike" is used, it isintended to include any and all rubber base products as well as allequivalent resilient products that could be used and which arecharacterized by a relatively low resistance to shear.

The basic mode of operation of the preferred configuration of our drillinvolves conventional rotation of the drill stem 10, FIGS. 2 and 3, inthe direction of the arrow 1, by applying sufficient torque to it at thesurface opening of the well, and, at the same time excitation of thevibrator 20 into axial reciprocating motion indicated by the doublearrow 2. This latter motion is imparted to the vibrator by a hydraulicactuator at the expense of some of the energy stored in the pressurizedfluid, this fluid normally being sent downhole for the purpose ofclearing the hole of the rock debris. Such a fluid may be natural gas,air, water,,oil, or drilling mud as is the common practice in rockdrilling. The hydraulic actuator is preferably of the piston type,comprising multiple flanges 35 and 41 formed into the vibrator 20 andhead 40 respectively, whereby fluid under pressure in alternate chambers34a, 34b, enclosed between 35 and 41, may exert an axial force as willbe described. The flanges 41 are indicated to be formed, for ease ofassembly, as part of an inner shell 40a within the outer shell 40b ofthe head 40. The inner shell 40a may be split lengthwise into two halvesto make assembly possible, and a threaded retainer ring 40c, with aconical load face to the bearing 12a, is used to hold the assemblytogether under preload. Other assembly means are possible.

Pressurized fluid is fed into the vibrator 20 by means of the inlet 21,is ported to the actuator by oscillatory means to be described, and isdischarged out of the vibrator through the outlet 22. The fluid thenenters a central hole 52 of the bit 50, or such other jets or courses asmay be provided in the bit. As explained in a later paragraph, shortlyafter the vibrator is set in motion, oscillatory power is transmitted tothe head 40 by means of the elastomeric bearings 12a, 12b, and by thehydraulic force generated by the pressurized fluid, provided thatcertain dynamic conditions are met; conditions involving the bearingresilience, the amounts of friction between moving parts, the magnitudeof the masses of the stem, head, and vibrator, and the crushing strengthof the particular rock. These conditions are investigated in a laterpart of this specification. The imparting of oscillatory power to thehead 40 causes it, and the attached bit 50, to vibrate axially in thedirection of the arrow 3, oppositely to the instantaneous direction ofthe vibrator mass 20, alternately compressing one of the bearings or 12bwhile simultaneously decompressing the other. On account of thisvibration an alternating pressure component is added to the averagepressure applied by the drill weight to the rock surface at the bottomof the hole. Thus, the rotary-vibratory character of the present drillis established by the combination of the vibratory movement 3 and therotary movement 1. Because the masses of the vibrator 20, head 40 andbit 50 are purposely made small by comparison with the mass of the stem10a and stem adapter 10, and because the elastomeric bearings 11 aredesigned to have the response of a soft spring to forced vibration, thestem is substantially prevented from participating in the oscillation ofthe rest of the members, thus preventing loss of energy and wear of thetopside rig parts and equipment.

The stem adapter 10 is the lower end member of the drill stem 10a, FIG.1, that is made up conventionally of sections of steel or aluminum pipeconnected end-to-end by means of a threaded end section, and has afourfold function; it carries the drilling fluid from the well surfaceto the bit 50 through a central channel 7; it forms the inner wall ofthe well annulus i that provides the return path for the drilling mudand rock debris; it helps provide static load on the bit necessary toeffect rock penetration; and it transmits the torque necessary to effectrotation of the drill as a whole supplied by the engine at the wellsurface.

It would be possible to use some other types of bearing spring seals12a, 12b than the laminated rubber-metal type preferred. For both thetop and bottom location, a nested stack of the springy thin dishes orarcuate washers known as Bellville springs could provide the highstiffness over a short range that is needed and simultaneously providethe needed alignment and sealing action and could be arranged totransrnit torque by means of keying or polygonal shaping. Thisalternative is illustrated in FIG. for the bottom location only, where astack of eight such Bellville springs are indicated as items 82.

Another possibility is to use a resilient elastomer with shear modulusso large that a direct use of it as an annular bearing spring sealwithout the need for interlaminations of metal could provide the neededhigh axial stiffness. In such a case, the angled load faces 13a, 13b,14a, 14b would still be desirable, producing a combination ofcompression and shear loading, and obviating the need for bonding of theelastomer against the shear load. A view is shown in FIG. 2b of thebottom bearing seal 83 only with large-angled load faces. Some bulgingof the elastomer may occur due to preloading plus relative displacement,and partial containment of it is provided at 84 and 85 by thecooperating shape of the parts 40 and 20 respectively. A polygonal shapeof the load faces may again be used to facilitate torque transfer.

The definition of a bearing or bearing spring seal as used hereinbetween the head 40 and vibrator 20 refers to an as sembly or group ofrelated parts that provides the functions of alignment, some axialcompliance, and a more or less complete sealing property, unless contextinfers otherwise.

An alternate method of providing bearing spring seal combinations isshown in FIG. 4a, in which a plurality of laminate bearings, rather thanonly a pair of bearings (12a, 12b as in FIGS. 2, 6), may be provided atany or all flanges 41 and 35 of the head 40 and vibrator 20respectively, in the annular gap between head and vibrator. That is, thebearings 12c, 12d, 12c, lZf take the place of the usual clearance gapexisting between the said flanges in FIGS. 2 and 6. This plurality ofbearings can provide sealing between each piston and cylindercombination of the fluid actuator, as well as aligning the head 40 andvibrator 20 and providing a stiff axial spring effect between them. Thishas the advantage of preventing wear between the coacting flanges andwalls due to an abrasive drilling mud, for instance.

Each bearing [20, 12d, l2e, 12f may be split to allow assembly, as maythe shells 40a, as previously discussed. The threaded retaining ring 40cholds the assembly together under preload. Each bearing may be made withsloped faces to provide high axial spring force as shown in FIG. 4a, orsome could be made with faces parallel to the longitudinal axis. Thebearing faces may also be made with circular or polygonal cross sectionsas previously discussed.

In FIG. 40, bearings 12c and 12d are sloped oppositely to bearings 12cand 12f, thereby providing for preloading of the one type against theother. Other combinations of sloping of the individual bearings areobvious, together with the split shells corresponding to 40a that willpermit assembly. Any of the several types of bearings discussed could beemployed.

Another variation of bearing and actuator means involving only onebearing would be provided in the case of a long rubber or laminatedelastomeric bearing sleeve interposed in the annulus between two longtelescoped sections corresponding to members of vibrator 20 and head 40.At one end of this elastomeric sleeve, an actuator could be formed by aseries of flanges alternately affixed to head and vibrator as alreadydescribed relative to FIG. 2. The clearance gap between the last(farthest from hearing sleeve) flange and coacting wall part of theactuator would then simply be open to the outside of the drill. That is,the entire configuration would be as in FIG. 2 except that bearing 12b,say, would be much longer axially, and not sloped, and bearing 12a wouldbe omitted.

I-Iaving indicated the basic mechanical configuration and mode ofoperation of our drill, we proceed now to describe the internalcomponents of the device and to develop a great depth of detailconcerning the operation of the various forms of the invention.

Clearly a very important component of the present rock drill is thevibrator 20 where the vibratory power is generated through aself-sustaining action. Its internal design involves either no movingsolid parts whatsoever, or a single moving member whose motion requiresno sliding of solid surfaces on one another. The two cases areillustrated in FIGS. 2, 3, 4 on the one hand and 6 on the other. The:nonmoving parts case, pictured in FIGS. 2, 3, and 4, is based on thealternating out put of a fluidic oscillator 23, which, in turn, is basedon the principle of momentum exchange explainable with the aid of FIG.5. A pressurized fluid, such as the drilling fluid in this case, isbrought into the source chamber 24 through the inlet opening 21 andproceeds into the interaction chamber 26 from which it starts streaminginto the output duct, say, 2712. Once this happens, the fluid jet stream28 attaches itself to the duct wall at 28b. However, part of the fluidflow through the duct 27b is diverted by the geometry of the device intothe feedback tube 30b and through the cavity 31b reaches the port 2%after some delay. This diverted flow entering the interaction chamber 26imparts momentum to the jet 28 at 28b perpendicularly to the directionof flow, and causes the jet 28 to switch over to the duct 270 where theprocess is repeated. Thus the flow of pressurized fluid oscillatesbetween the ducts 27a and 27b with a frequency set by the mechanicalresistance and length of the feedback tubes 30a, 30b, and by themechanical capacitance of the cavities 31a, 31b. Additional feedbackdelay could be obtained if needed by cascading additional fluidicamplifiers in the return paths. As a matter of fact, the two flows outof 27a, 27b may be made of different durations by making the twofeedback paths (30a, 31a) and (30b, 31b) different in shape and/or sizefrom each other. The oscillators geometric configuration shown in FIG. 5is one of a great number of possible configurations, and is used hereonly by the way of explanation of the fluidic oscillator in general.Accordingly, the term fluidic oscillator as used herein means a fluidicdevice whose output flow is switched automatically, and without theinterference of moving solid parts, other than the vibrator 20 and head40, between two output ducts by means of a feedback path for each duct,that accepts part of the output flow or energy at some downstream pointof the output duct (the duct end point included) and delivers it,delayed, at a point upstream at the interaction chamber 26.

Returning now to FIG. 3 we observe that the fluidic oscillator 23 isformed inside the solid mass of the vibrator 20. In practice thevibrator 20 may be cast in two parts, each of a half-cylinder shape, forthe easy forming of the oscillators cavities and channels, and the twoparts welded or otherwise connected together.

The oscillators duct 27a communicates with the annular piston chamber340 by means of the branches 320, while the duct 27b communicates withthe annular piston chamber 34b by means of the branches 32b. Thechambers 34a. 34b are formed by the flange 35 of the outer surface ofthe vibrator 20 and by the flanges 41 of the inner surface of the head40. When pressurized fluid from the oscillator acts upon the chamber340, it tends to force the vibrator 20 downward and the head 40 upward;by contrast, when pressurized fluid acts upon the chamber 34b, thedirections of these forces are reversed. The clearance gaps 36 are madewide enough that fluid leakage out of the chambers 34a, 34b through themis allowed. The fluid to mechanical energy transfer efficiency isthereby reduced but some leakage path is nevertheless necessary; bothfor purging the respective chambers 34a, 34b during reversal and becausewithout it the oscillator will become choked and the switching of theflow between its two ducts 27a and 27b impossible. This leakage fluid issiphoned through the passages 37 into the outlet opening 22 and formsthe main flow of drilling fluid descending into the drill bit anddischarging through the bit passages. Although only one pair of pistonchambers is shown in the drawing, it should be obvious that more suchpairs may be employed in a parallel battery along the length of thevibrator, and so may other fluidic oscillators be used in parallelarranged to act in concert.

In the design variation shown in FIG. 4 the fluidic oscillator feedbackpaths 30a and 30b emanate from the piston chambers 34a and 34brespectively. Thus the feedback action is coupled to the relative motionbetween vibrator and head 40. Since, in turn, the latter motion isstrongly influenced by the crushing strength of the particular rockencountered by the bit, the oscillator's switching becomes dependentupon the rock characteristics. This is a desirable feature in that itmakes the oscillation continuously self-adjusting as the nature of thepenetrated strata changes. The starting points a, 25b of the feedbackpaths a and 30b can be either in close proximity to or away from thecorresponding duct 27a, 27b end points, or even placed somewhere in thegap 36 between vibrator and head; in the latter position the feedbackprocess may include a judicious valving by the movement of flange 41. Inaddition to this variance in the fluidic oscillator paths FIG. 4 alsoillustrates a different usage of the discharge passages 37. The latterare now placed on the walls of the head and discharged the fluid spentin the chambers 34a, 34b directly into the wells annulus where lowerpressure of the drill fluid prevails. In this case, however, a conduit60, originating at inlet opening 21, has to be added through theactuator 20 to supply the flow necessary at the cutting surfaces of thebit. A restricted open ing or nozzle is provided within the bit body,directed in a conventional manner so that substantial part of thehydraulic pressure of the fluid is converted into velocity through whichcleansing of the bit and removal of the rock debris from the well bottomis effected. This restriction is such that necessary pressure foroperation of the fluidic oscillator is maintained. The actuatorembodiment shown in FIG. 6, like that of FIGS. 2, 3, and 4, features thebattery of piston chambers 34a, 34b formed between vibrator 20 and head40 and fed respectively by the passage pairs 37a, 37b. However, theplace of the fluidic oscillator is now taken by the pipe-shapedmechanical oscillator carried within the main conduit or tubular cavity6 of the vibrator 20 preferably by means of the elastomeric bearings,springs and seals 15 and caused to oscillate axially relative to thevibrator 20 through a small distance shown by arrows 4 by means laterdescribed. Other support and sealing means are possible, as by two plainbearings, with springs arranged to center the oscillator 60 when atrest. Such a plain bearing could also perform the sealing functionrequired. Here, a support member is defined as having associated sealingand centering properties. Jet nozzles 61 through the lateral wall of theoscillator face the passage pairs 37a, 37b over a small clearance gap,the oscillator itself being always filled by the pressurized fluid.Depending on the relative axial position of the vibrator 20 andoscillator 60, and because the inner openings 38a and 38b of thepassages 37a, 37b respectively are offset axially with respect to oneanother, pressurized drill fluid from the jet nozzles 61 is forced intoone or the other piston chambers 34a, 34b alternately while the spendfluid exits simultaneously from the other. finally leaving outlet port22 via ports 25d. This alternate action forces the head 40 and vibrator20 to reciprocate in opposite directions as in the embodiment shown inFIG. 3. It should be noted that these reciprocal motions along with thearrangement of the jet nozzles 61 opposite the openings 38a, 38b resultin automatic valving of the pressurized fluid into the chambers 3411,34b. and that the sense of the generated forces is such as to causeselfsustaining oscillation.

In FIG. 6, large ports 250 are provided through the vibrator wall toallow pressure equalization in the chambers next to the bearings 12a and1212. It would be possible instead to omit the top and bottom flanges41, which would result in the bearings 12a and 12b respectivelythemselves forming more or less functional flanges against whichpressure forces may act. However, since the bearings 12a, 12b are joinedto the vibrator 20 at their inside radii, some of the force transferredto them will be transferred to the vibrator rather than all to the head40, rendering said bearings less effective than the flange 41 for forceproduction. However, if by design the outside bearing faces and matingload surfaces of head 40 at 14a, 14b were placed closer to the centrallongitudinal axis than is the case in FIG. 6, then more effectiveflanges would result, fuctionally. Thus, any piston chamber, as in FIG.4a, for instance, may be defined to be more or less effective for forceproduction as the functional flange and piston flanking it are effectivein that sense.

Restricted outlet paths 25h, FIG. 6, may be provided to maintain fluidcirculation and prevent it from stagnating and possibly forming soliddeposits within piston chambers.

The end plug and port device 62 has the function of separating the highpressure region within oscillator 60 from the relatively low pressureregion outside the oscillator and in the port 22, while still providingaccess between the latter two through the port 25d. A passage 64 isnecessary as a connection between the high pressure area and theenclosed volume 65 to equalize pressure on the top and bottom of themechanical oscillator 60 and bearings 15. Otherwise, a considerabledownward steady load on the oscillator 60 could occur. In the methodshown, the plug 62 is separated from the oscillator 60 by a clearancegap 63 (drawn exaggerated) that is made as narrow as possible to preventflow loss.

In some cases of design, it may be desirable to have a means ofproviding an additional damping effect on the motion of the mechanicaloscillator 60 beyond what would otherwise exist, i.e., provide a forceopposing its velocity. This can affect its relative phase of oscillationwith respect to the vibrator 20, thus affecting the efficiency andextent of fluid power transfer. One means of providing such damping isreadily available by making use of the enclosed volume 65 and passage64, which can be made small enough to result in significant pressuredrop across its length as the fluid is forced back and forth through itdue to the piston action of the bearing 15 and bottom of the mechanicaloscillator 60.

Another means of blocking flow from the bottom of mechanical oscillator60 and providing an exit port for spent fluid is shown in FIG. 60, inwhich high pressure fluid has direct access to both ends of oscillator60, and spent fluid is allowed to exit through the ports 25c into aspace above the bottom bearing 12b and then out through the exit port67.

There are shown in FIGS. 6a and 6b several other ways of arranging thejet porting corresponding to the typical view within the circled area ofFIG. 6. In both of these, there are two jet nozzles 61a and 61b, to takethe place of the single nozzle 61 at every location on FIG. 6. Nozzle61a separately discharges only into passage 37a, while nozzle 61bdischarges only into passage 37b in both FIGS. 6a and 6b, wherein theoscillator 60 is shown instantaneouly positioned so that nozzle 61b isaligned with its passage 37b, while nozzle 61a is cut off. Thedifference is that the flow passages in FIG. 6b are not crossed,necessitating a relative phasing of motion of oscillator 60 that differsby l from that in FIGS. 6 and 6a. In both new cases, however, thepassages and their respective nozzles are parallel, necessitating lessdirectional change of the fluid, reducing wear on the passages in thecase of some drilling muds which may be abrasive.

Other variations in porting arrangements may be visualized withoutdifficulty.

The initiation of relative oscillation and valving by oscillator 60depends upon the ever present small turbulance in the flowing drillfluid and upon the mechanical vibration of the bit moving over theuneven rock face, both these factors being reflected in a jitter typemovement of the mechanical oscillator 60. Once this jitter causes thefirst valving and'the vibrator 20 is forced into its initial strokes,the oscillator 60 by design does not follow the actuators motion. Thiseffect is achieved by making the natural frequency of the oscillator 60very low compared to the relatively high natural frequency of thevibrator. To satisfy this condition the equivalent spring constant ofthe elastomeric bearings is made relatively low by comparison with theequivalent spring constant of the laminate bearings 12a and 12b.Although the first few strokes may be rather weak resulting in onlypartial valving, the energy imparted by them to the vibrator forces itinto a movement of larger and larger amplitude and, therefore, into moreand more complete valving and the concomitant actuator energy buildup,with the head and bit masses finally involved in the oscillation as inthe previous case.

p A combination of the already explained fluidic and mechanical meansfor inducing and/or maintaining the oscillatory motion of the bit isillustrated in FIG. 7. As seen there, the arrangement of the mechanicaloscillator 60, suspended by means of the elastomeric bearings 15 withinthe vibrator 20, is identical with the arrangement already discussed inconnection with FIG. 6. However, the mechanical oscillator 60 is setinto a reciprocating motion by an hydraulic actuator, as by a source ofoscillating fluid acting on a piston, preferably by a small auxiliaryfluidic oscillator 70 which is formed within the body of the sleeve 77inserted and rigidly affixed within the body of the vibrator 20. Thesource chamber 71 of this auxiliary oscillator is fed the pressurizeddrilling fluid through the auxiliary inlet 72. The output ducts 73a and73b alternate in supplying the annular piston chambers 74a and 74b withthe pressurized fluid which acts alternately upon one side or the otherof the flange or piston 75 causing a reciprocating motion of themechanical oscillator 60. The drain channels 76 serve to carry the spentfluid away from the piston chambers 74a, 74b into the low pressure space78 between oscillator 60 and vibrator 20, and thence eventually into thewell hole. Feedback paths 80a, 80a, with capacitance cavities 81a, 8111,or other means of delaying the feedback signal if needed, are providedbetween the reaction chamber 79 of the fluidic oscillator and each ofthe output ducts 73a, 73b. Obviously more piston chambers acting inparallel with those described are possible. The fluidic oscillator driveis provided here to positively establish relative motion of themechanical oscillator 60 and vibrator 20 at about the resonant frequencyof the vibrator itself, so that the startup of vibrator oscillation doesnot have to depend upon chance vibration causing the start ofoscillatory valving. Such startup with the previous simplerconfiguration of FIG. 6 would probably become increasingly moredifficult as the bearing springs 12a, 12b are designed to be stiffer forhigher resonant actuator frequency. Addi' tionally, the presentconfiguration of FIG. 7 could maintain higher amplitudes of relativeoscillation between oscillator 60 and vibrator 20, providing bettervalving while allowing smaller oscillation amplitudes between vibrator20 and head 40 and the latter with respect to the rock.

It is indicated with respect to FIG. 7 that the forced oscillation ofthemechanical oscillator 60 is parallel to the drills lon' gitudinal axis.However, the functions indicated could be served as well by a torsionaloscillation instead, with the jet nozzles 61 or 61 and 61b and receivingpassage pairs 3711, 37/2 suitably respositioned for coaction torsionallyover a small angle. Conical laminate bearings 15 would suffice to allowsuch motion while preventing axial motion, and a rotational vane typefluid actuator would suffice for oscillatory torques in place ofthepiston actuator shown.

The auxiliary fluidic oscillator 70 could be designed to be primarilyeffective in starting up the vibratory motion of the system rather thanalso being responsible for maintaining it. That is, when the growingamplitude of vibration of the vibrator 20 reaches a certain point, thespring or inertial forces applied to the mechanical oscillator 60 inFIG. 7 due to the vibration (e.g., through the bearing or supportsprings 15) could become adequate to continue its oscillation in themanner already described with respect to FIG. 6. This could result ifthe system is designed so that actuation forces on the piston 75 of FIG.7 become relatively weak under this condition compared to the otherforces on the mechanical oscillator 60, including support spring andinertial forces. This could be a consequence of the sizing of thefluidic oscillator 70 and its associated piston 75. The character of thesteady state vibration of the mechanical oscillator 60 and that of thewhole system could still be influenced beneficially by the continuedaction of the forces of piston 75, however; e.g., by providing a dampingaction.

Deviations in the disclosed connections of the fluidic oscillator 70 andpiston 75 are believed to he obviously within the spirit of theinvention, and furthermore, different kinds of fluid oscillators andactuators to perform the same task are intended to be included. Forinstance, a reaction type actuator or an unbalanced rotating weightconnected to the mechanical oscillator 60 could be used to drive it, aswill be discussed further at a later point.

In FIG. 6 the mechanical oscillator 60 is described to be positioned atevery instant with respect to vibrator 20 by means of its inertiacausing it to be effectively stablized. However, the phase relationshipestablished thereby between the mechanical oscillator and the vibratormay not be optimal.

Another means of relatively positioning the mechanical oscillator 60 isto provide force-transmitting means to it not only from the vibrator 20but also from the head 40 to the mechanical oscillator 60. A relativephasing relationship wherein the position of the mechanical oscillator60 is lagged may be provided, for instance, where the mechanicaloscillator may undergo a substantially cosinusoidal oscillation ofposition while that of the vibrator is sinusoidal and that of the headis negatively sinusoidal. This can be done by connecting the mechanicaloscillator 60 by an axial spring to one of the massive members (head 40or vibrator 20) and by a damper to the other (vibrator 20 or head 40).In FIG. 6d, the damper connection is to the head 40, via damping piston75, while bearing springs 15 provide the spring connection to thevibrator 20 as before. In opposite fashion, in FIG. 62, the damperconnection is made to the vibrator 20 via damping piston 75, while thebearing spring 15b, attached between the head 40 and mechanicaloscillator 60, has much greater axial stiffness than the bearing 15a, sothat the behavior effectively is determined by the former bearing spring15b.

In FIG. 6d, the multiple passages 25c and 67 provide a way for spentfluid to exit from the annular space surrounding the mechanicaloscillator 60, while in 62, passages 67 do the same.

In FIG. 6d the passages 65a provide a means whereby the space 65 may beequalized in pressure with the interior of the mechanical oscillator 60.The passage 65b provides similar equalization of pressure for the space68, whereby average pressure loads on the mechanical oscillator 60 maybe made negligible. The damper piston chambers on either side of piston75 have restricted outlets formed, for instance, by appropriately-sizedgaps 63a, 63b at flanges 75b, 75c and/or by a bypass formed by a gap 63between the piston 75 and its cylinder wall. When the mechanicaloscillator 60 is moved axially relative to the head 40, a pressure dropdepending upon relative velocity will be created forcing flow throughthe gaps mentioned and creating a damping force equal to the pressuredrop times the piston area. Multiple damping pistons and flanges mayobviously be provided to increase the damping force if desired, in themanner of the main driving actuator. Restricted orifices 64 may beprovided to allow a constant flow from the damper piston on either sideof piston 75 to the low pressure region via passages 67 in order toprevent fluid stagnation and solidification in the damper pistonchambers. If enough flow is so allowed it would be possible to obtainmore constant damping characteristics for various oscillatoryamplitudes, since the average damping flow would be shifted from zerowhere the damping characteristics change most rapidly. It is desirablethat the mass of mechanical oscillator 60 be made relatively small sothat its position is determined mostly by the coaction of the springs 15and damper.

In Fig. 6e, although the bottom bearing 15b is attached to head 40 andis represented to have high axial stiffness, the bottom bearing supportof the mechanical oscillator could as well be attached to the vibrator20 as in FIG. 6, with an axiallyoperative spring attached between thehead 40 and the mechanical oscillator 60.

Also in FIG. 62, the damper piston chambers on either side of piston 75have restricted outlets formed, for instance, by an appropriately sizedgap 63a at the flange 75b and holes 63b,

both outlets communicating with the high pressure region. A

bypass gap could also be provided at 63 as in the former case. fRestricted orifices 64 could be provided for the same purpose asdiscussed vis-a-vis FIG. 6d. Passages 25c allow equalization of pressurein the space above top flange 41 at the low pressure in the annulussurrounding the mechanical oscillator 60. The mass of the latter isagain made relatively small as before.

It would otherwise be possible for the damping flow to occur primarilythrough the restricted orifices 64 rather than through the gap and holes63a, 63b, or bypass 63, so that the latter could be omitted. It willthen be noticed that these passages 64 correspond to the passages 76 ofFIG. 7.

It appears obvious that an auxiliary fluidic oscillator could beprovided and joined to the piston chambers as in FIG. 7, with thepurpose of providing a starting means. When the mechanical oscillatormotion amplitude builds up to a certain point under the influence of thefluidic oscillator-driven mechanical oscillator 60, the increasingdamper and spring lSb forces could cause a shift toward the foregoingtype of operation in the steady state, wherein the mechanical oscillatormotion is primarily controlled by the head 40 and vibrator 20 motionsthrough spring b and the damper respectively.

From the description of the invention presented so far, it is obviousthat the function of the drill is based upon the dynam ic interactionbetween the vibrator 20, on the one hand, and the head 40 and bit 50 onthe other. This interaction, which leads to the desired transfer ofmechanical energy from the vibrator to the bit and eventually to therock, is controlled by the magnitude of the two masses of vibrator andhead-plus-bit, by the strength of the spring effect of the laminateelastomeric or other type bearings 12a, 12b connecting these two masses,and by the crushing strength of the particular rock on which the bit isworking at a given time. These four mechanical parameters make the drilla system of two oscillators coupled to each other and to the rock and,therefore, a system possessing a natural frequency of oscillation.Consequently, as the nature of the rbck changes during the drillingoperation so may the natural frequency at a given amplitude change,although the amount of the latter change may not be large. Generally, aninherently oscillatory system operates at maximum amplitude when thefrequency of the force actuating it is substantially equal to thesystem's natural frequency. Fluidic oscillators, like the one used inthe present invention to generate this force can be induced to changetheir natural frequency by various means, one of them being the fluidspressure. We wish, however, to show at this point one particulartechnique for controlling the natural frequency of the system optimallywithout outside intervention from the well's surface. The technique isillustrated in FIG. 8 and is based on the use of the acceleration forceof the vibrator for imparting momentum to the fluid stream in thecavities of the fluidic oscillator 23 and in a manner effecting theswitching of the flow between the ducts 27a and 27b alternately. Asshown in FIG. 8 the oscillator 23 is placed transversely within the bodyof the vibrator 20 with the plane of its feedback paths 30a, 30bparallel to the vibrators longitudinal axis. The source chamber 24 isconnected to the main inlet opening 21 through inlet 19. During thefirst few moments of the oscillators operation its flow is switchedbetween 270 and 27b by the action of the feedback paths 30a and 30b asalready described. However, after a number of vibrator strokes at ornear the natural frequency of the resonant system, sufficientacceleration is imparted to the vibrator 20 to beneficially affect thefunction of the oscillator. Specifically, if at a given instant thefluid is streaming through the duct 27a into the piston chamber 34a andthe vibrator is moving downward, an upward acceleration is applied to itby the bearing springs I211, 12b as it nears the end of the downwardstroke; consequently, an inertial force is applied to the stream at theinteraction chamber 26 forcing the stream to switch over to the duct27b. This brings the pressurized fluid into the piston chamber 34b, thusreversing the force applied on the vibrator 20 and urging it upward. Atthe end of the upward stroke, the inertial force appearing on the fluidin 26 is pointed upward pushing the stream upward and causing a newswitching, and this process will continue ad infinitum. While the duct27b is operative and fluid begins to flow through the correspondingfeedback path 30b, the inertial force acting at 26 is seen to alsoreinforce the feedback flow through 29b thus adding to the initial causeof switching the main flow from 2712 to 27a. A similar statement may bemade when 27a is operative. Since the oscillator flow switching due toacceleration forces is directly dependent upon the motion of thevibrator 20 which in turn is determined by the crushing strength of therock along with the mechanical parameters of the drill system itself,the system sets its frequency of oscillation automatically and inaccordance with the prevailing physical condition of the rock formationthat is being drilled. This effect we choose to call motional feedbackfrom the vibrator to the oscillator. While only one combination of afluidic oscillator 23 and a pair of piston chambers 34a, 34b are shownin FIG. 8, it should be obvious that two or more such combinations maybe used within the same actuator 20.

The same principle of motional feedback may be applied to the case ofthe fluidic oscillator 70 and vibrator 20 of FIG. 7, in which theauxiliary fluidic oscillator 70 is already indicated to be horizontal.In this case, because of the inclusion of an additional component takingpart in the oscillation of the overall system (Le. the oscillator 60)different phasing problems may be encountered from the case of FIG. 8.Phasing lag or lead mechanisms may need to be incorporated into thefluid paths in order to make the feedback phasing correct forself-oscillation. Motional feedback switching may be feasible withsmaller amplitudes of oscillation in this case because of the smallerdimensions of the auxiliary fluidic oscillator 70.

It is apparent that driving of the mechanical oscillator 60 in the modedescribed in connection with FIG. 6 involves an imparting of force to it(60) that causes it to oscillate and switch flow to piston chambers 34aor 34b. This imparted force depends upon the motion of the vibrator 20,so that the previ ously-discussed operation of the mechanical oscillator60 in the system can also be defined as an instance of motionalfeedback. Other mechanizations or variations of the principle appearobvious.

From the description of the vibrator 20 given so far, it can be seenthat the two primary preferred versions of the device relative toobtaining oscillating flow at the piston areas feature either no movingsolid parts within the vibrator, or only one such part, and that allmotions between vibrator and drill stem or between vibrator and headinvolve no sliding of solid surfaces against each other. Instead,advantage is taken of the excellent charcteristics of the preferredelastomeric bearings as far as mechanical strength, wear, resilience,and sealing are concerned to accomplish the desired task.

Some other fluid actuator means of providing an oscillating force tocause vibration of the vibrator and head-bit system are possible also.Specifically, as seen in FIG. 9, one way is to direct the output ducts27a and 27b of the fluidic oscillator 23 to nozzles 22a and 22 directedupwardly and downwardly into the drill hole space, causing alternateupward and downward reaction forces to be exerted upon the vibrator 20.

Another actuating technique is to mount unbalanced rotating weights,driven by fluid turbine means (or by any rotational energy actuator,such as an electric motor), to the vibrator 20, the centrifugal forcesof the unbalanced rotating weights causing an axial component ofoscillating force on vibrator 20. In the previous art, this has beendone with long solid vibrating columns, but in the present device withtwo discrete masses (vibrator and head-bit) separated by novel bearingspring means as described, a more compact vibrating arrangement can beprovided. Contrarotating pairs of equal offset weights geared lzl arepreferable, to cancel or minimize lateral forces.

Either of the above two techniques could be used for providing actuationforces on the mechanical oscillator 60 as previously mentioned.

In all variations of the subject rock drill discussed so far, the stemadapter has been connected to the vibrator preferably by means of thelaminate bearings 11. However, in all these cases the drill can be madeto function equally well if the stem adapter 10 is connected to the head40 by means of the relatively soft elastomeric bearings 11 (preferablylaminated), while the vibrator is connected only to the head 40 throughthe relatively hard bearing springs 12 as shown in FIG. 10.

Another variation in design is illustrated in FIG. 11, in which the bit50 is attached to the inner massive member 20 of the drill and the stemadapter 10 is connected to the outer massive member 40, a reversal ofthe cases described to this point. This rearrangement of connections tothe vibratory massive members 20 and 40 results in essentially the sameinternal features as heretofore discussed except that phasing and flowswitching parameter revisions may be required in th cases of motionalfeedback.

In the embodiments of the invention presented so far only thearrangements causing the axial reciprocating motion 2, FIGS. 2, 3, 4,and 6 were explained. At this point, however, we wish to discloseadditional means for implementing the rotary motion 1 as well. Thesemeans comprise a hydraulic turbine wheel 42 shown in FIGS. 12 and 12a toconsist of a plurality of radially placed turbine blades 43 rigidlyattached to the interior wall of the head 40, and of a plurality offeeder ducts 44 provided in the body of the vibrator 20. The feederducts 44 are connected to a central conduit (not shown), similar to 6aof FIG. 4, and direct high velocity jets of pressurized drill fluid ontothe blades 43, imparting a substantial portion of the mo mentum of theirfluid to the blades and, therefore, applying a torque to the head 40.This torque forces the head 40, and bit, to turn on the conventionalbearings 39. Although only one turbine wheel 42 is shown in the drawingsit should be obvious that a battery of such wheels may be used as wellifthe torque requirements dictate it. Thus the need to turn the wholedrill assembly from the surface of the well is eliminated as is the needfor the drill stem 10 to extend all the way to the top of the well. Thisis so because the feeder ducts 44 are directed along radii of thevibrator 20 so that the reaction force of the outgoing jets generates noreaction torque on the vibrator 20. Thus, instead of the rigid drillstem, a flexible stem or power hose 46 connecting the pressurized fluidsupply to the inlet 21 through a rotating joint 45 suffices to supplythe necessary power to the whole drill, and results in a substantialsimplification and economy of the surface equipment. It will beappreciated that the power hose is not required to transmit torque and,therefore, can be made sufficiently flexible to be stored in spools.Furthermore, this embodiment materially reduces torque requirementssince there is no rotating lengthy drill stem to be handled. Afterimpinging on the blades 43 and delivering energy to them, the drillfluid is diverted into the cavity 47 in a vortexlike fashion andeventually discharged from the bit passages onto the bottom of the well.In this case the length of the drill stem is only sufficient to providethe static weight required for pressing the bit against the rock forefficient drill action and to control hole deviation from the vertical.Hence the drill stem may comprise only the stem adapter 10.

A modification ofthe means for implementing the aforesaid rotary motion1 further reduces the torque requirements. The modification consists ofthe feeder ducts 44 being connected to the branches 32a of theaforementioned fluidic oscillator 23 instead of the ducts beingconnected to the inlet 21 through a conduit. This causes fluid jets toissue only when the annular piston chamber 34a is active and, therefore,torque to be applied to the head 40 at the time the latter executes itsupward stroke partially relieving the force between the bit and the bot=tom of the wellv This, of course, is the most appropriate time to indexthe bit since the coulomb friction between bit and rock is at a minimumresulting in a reduction of bit wear.

An alternate design for effecting bit rotation through use of mechanicalenergy released downhole is the configuration shown in FIG. 13 whereinthe laminae of the elastomeric bearing 12b are stacked into sections ofparallel helical surfaces, instead of being parallel conical surfaceswhich is the case with the bearing 12a, these sections coacting with thehelical faces 12h of the section 13b of the vibrator. Such a bearingwhen compressed in the axial direction undergoes a deformation parallelto the helical surfaces; this deformation possesses both an axialcomponent and a rotary component of which the former is made to providethe vibratory action against the rock bottom of the well, while thelatter will cause rotary displacements of the vibrator and head inopposite directions; the mechanism involved is as follows: in the steadystate of the vibratory motion the head 40 and vibrator 20 are moving inopposite directions; therefore, during a down stroke 2dr ofthe latterand upstroke Zuh of the former the compression of the bearing 12b isaccompanied by a rotation ofthc vibrator to the left. llv and of thehead to the IIght I f/l. Since the head and bit is pressed against therock at all times the extent of lr/i in degrees of arc will be smallerthan the extent of llv; i'.e., lIv=u lrlz=h, and a b,. When the strokesare reversed and the vibrator moves upward, Zur. While the head movesdownward, 2(1/1, the bearing deformation will be removed since thecompression is relieved. In the course of doing so the vibrator willrotate to the right, lrv=a,, and the head to the left, l/h=b,. with u bfor the reason given above. The net result after one oscillation will bean angular displacement u,u,. for the vibrator and Iii-l1, for the head.Since the bearing 12b should be left with no additional stress at theend of any one oscillation than at the beginning, we should have uu,.=l/,.. But during the heads downstroke the increased pressure againstthe rock makes 11f u,f': when this inequality is substituted in the lastequation we obtain /i,." b Thesetwoincqualities mean that there is a netrotation of the head 40 to the left and of the vibrator 20 to the right;the former results in bit indexing, while the latter is accommodated bythe rotary joint 45 between vibrator and stem adapter 10.

In the two types of self-rotating rock drill just discussed (FIGS. l2,13) the stem adapter 10 had no direct connection to the surface. Anotherapproach is to have it attached to a cable from the surface. As shown inFIG. 14, the eyelets 17 rigidly attached to the upper end of the: sternadapter 10 serve for fastening a lift cable through which the weight ofthe drill is partially counteracted by pulling on the cable as drillingconditions require. The elastomeric bearing ll, besides attenuating theupward transmission of vibration, permits drill weight reduction withoutlifting the bit off the wells bottom. Means of attaching the cable tothe stem adapter 10 other than eyelets may be used. In all otherrespects the drill assembly remains the same as disclosed previously,with the vibrator 20 being resiliently attached to the head 40 by meansof laminate elastomeric bearings 12a and 12b, and with externalattachment of the vibrator to the stem adapter 10 as discussed earlier.The respective external-internal relationships of vibrator and adaptercould be reversed as well.

The dynamics of the three-mass system of FIGS. 2, 3 and 4 is modeled inthe mechanical circuit diagram of FIG. 15 that in turn leads to thefollowing system of three differential eq uations in the unknown timevarying displacements x=x(r). y= yll). 2 20):

In this system M M M are the masses of the stem 10, vibrator 20, andhead-bit 40, 50; W W W;, are the corresponding weights; (k k and (n, rare the spring constants of the elastomeric bearings 11 and 12respectively, and

- the viscous friction coefficients associated with the motion of I Mand M f the reciprocating force supplied by the oscillator 23; and x, y,z the displacements of the stem, vibrator, and bit from the bottom ofthe hole in that order; finally r is the friction coefficient betweenhead-bit and the sides of the hole in all probability nonviscous.

The alternate arrangement of FIG. 10, whose equivalent mechanicalcircuit is seen in FIG. 16, is correspondingly described by the systemof equations in a similar fashion the four-mass system of FIG. 6 ismodeled by the mechanical circuit diagram of FIG. 17 from which thefollowing system of differential equations is formed:

in this system M is the mass of the mechanical oscillator 60 and A1,,in, the spring constant of the laminate bearings 15 and the coefficientof the viscous friction encountered by M In all three systems Wsrepresent the corresponding weights and is the force-displacement curveassociated with the crushing strength of the particular rock. Typically,C(z) has the form shown in FIG. 19 with the curve parameters dependentupon the physical characteristics of the rock and upon the geometry ofthe drill bit. Although the term G(z) is nonlinear, solution andoptimization of the above systems is possible using eithernonlinear-analytical or computer-based techniques.

Finally, the system of FIG. 14 is represented by the circuit diagram ofFIG. 18 from which we obtain:

in all these systems the presence of the terms containing the quantity(z-y) in the equations for M a y/dt and M d"z/dt explains and describesquantitatively the interaction between vibrator and head-and-bit. Theterm G(z) on the other hand, describes the mechanism of the transfer ofmechanical energy from the drill to the rock.

While full and complete disclosure of the invention has been set forthin accordance with the dictates of the patent statutes, it is to beunderstood that the invention is not intended to be so limited. It willbe apparent to those skilled in the art that various changes may be madeto the embodiments described herein without departing from the spirit ofthe invention or the scope of the appended claims. Moreover, it isapparent that the vibratory system resulting could be used inabove-ground applications as for tamping or pile-driving purposes, or asa vibration generator in conveyors or other machinery. In this case, theend connections in lieu of stem and bit may be adapted to the particularpurposes at hand, so that the words bit" and stem or stem adapter may bebroadly construed. Also, it is not necessary that the connecting memberstaking the place of the bit and stem adapter necessarily be on oppositeends of the assembly. Yet another important application of our vibratorysystem is in the manually held air hammer where noise and operatorfatigue can be substantially lessened. Still other applications includethe implementation of vibratory drilling of hard metals or 5 other typesof material removal as in mold-carving, dental work, etc.

We claim:

1. A vibratory drill assembly as described including:

an inner massive member at least partially telescoped within an outermassive member, the two said massive members being separated by, andheld substantially concentric to each other about a common longitudinalaxis by at least one bearing member having sealing properties and stiffaxial spring characteristics whereby limited axial displacement of theone massive member with respect to the other results upon theapplication of axial force between 4 said massive members;

a bit drivingly attached to one of said massive members and a stemadapter attached resiliently in said axial direction to one of saidmassive members; and

a fluid-operated actuator producing force between said two massivemembers, including at least one piston functionally formed on the outersurface of the teleseoped portion of said inner massive member andcoacting with the inner surface of said outer massive member, and atleast one flange functionally formed on the inner surface of said outermassive member and coacting with the outer surface of said inner massivemember to provide at least one effective piston chamber to acceptpressurized flow cyclically through porting means from oscillatoryconverting rneans within said inner massive member, said oscillatoryconverting means acting upon an inlet flow of pressurized fluid suppliedthrough said stem adapter via an internal inlet passage and convertingsaid inlet flow to said cyclic flow, whereby fluid energy may beconverted to vibratory energy carried by the resonant system formed bysaid inner and outer massive members and said springlike behavingbearing members coupling them together, said vibratory energy dissipatedby said bit repeatedly striking the material to be penetrated by saiddrill assembly.

2. The device of claim 1 wherein each one of at least two said bearingmembers is of the laminated elastomer-metal type as described and hasload faces inclined with respect to said longitudinal axis, said bearingmembers being preloaded against one another.

3. The device of claim 1 wherein each one of at least two said bearingmembers is make of a stiff resilient elastomeric material as describedand has load faces inclined with respect to said longitudinal axis, saidbearing members being preloaded against one another.

4. The device of claim 1 wherein each one of at least two said bearingmembers is composed of a nested stock of thin springy dish-shaped metalwashers as described, said bearing members being preloaded against oneanother.

5. The device of claim 1 wherein at least one of said bearing membershas a noncircular cross section in a plane perpendicular to saidlongitudinal axis, whereby said bearing member may transmit torquewithout substantial windup.

6. The device of claim 1 wherein said bearing members are two in number,forming an opposed pair on opposite ends of said fluid actuator andsealing same, preloaded against one another.

7. The device of claim 6 wherein said bearing members are laminateelastomer-metal bearings as described and have load faces inclined withrespect to said longitudinal axis, being formed of joined flat sectionsso as to have a polygonal cross section in a plane perpendicular to saidlongitudinal axis, whereby said bearing members may transmit torquewithout substantial windup.

8. The device of claim 1 wherein said bearing members are interposed inthe annular spaces between said pistons and said coacting inner surfaceof said outer massive member and between said flanges and said coactingouter surface of said inner massive member, acting to seal said pistonchambers from one another.

9. The device of claim I wherein said stem adapter is engagedtelescopically to one of said two massive members, said internal inletpassage comprising longitudinal connecting holes within said stemadapter and said massive member, at least one relativelyaxially-resilient elastomer-containing sleevelike member interposed inthe annulus between said telescoping members, connecting and aligningthem, sealing between them against said inlet flow of pressurized fluid,and transferring axial load force from said stern adapter to saidmassive member, whereby limited axial movement of said massive member ispossible with respect to said stem adapter through shear strain of saidelastomer, resulting in effective isolation of said stem adapter fromsaid resonant system.

10, The device of claim 9 wherein each said elastomer-containing memberis a laminated elastomer-metal bearing as described.

H. The device of claim 9 wherein mechanical stop means are provided toprevent axial overstressing of said elastomercontaining sleevelikcmember.

12. The device of claim 9 wherein at least one said elastomercontainingmember has a noncircular cross section in a plane perpendicular to saidlongitudinal axis, whereby said bearing member may transmit torquewithout substantial windup.

13. The device ofclaim 9 wherein each said elastomer-containing memberis a laminate elastomer-metal bearing as described, being formed ofjoined flat sections so as to have a polygonal cross section in a planeperpendicular to said longitudinal axis, whereby said bearing member maytransmit torque without substantial windup.

14. The device of claim 13 wherein said stem adapter is telescopedexternally to said inner massive member, and said bit is attached tosaid outer massive member on the opposite end of said drill.

15. The device of claim 13 wherein said stem adapter is telescopedinternally to said outer massive member and said bit is attached to saidouter massive member on the opposite end ofsaid drill.

16. The device of claim 9 wherein said stem adapter is telescopedexternally to said inper massive member.

17. The device of claim 9 wherein said stem adapter is telescopedinternally to said outer massive member.

18. The device of claim 1 wherein said oscillatory converting meansincludes at least one fluidic oscillator, said fluidic oscillatorsalternately feeding said inlet flow of pressurized fluid into alternatepiston chambers respectively through one of a pair of output ducts, eachsaid fluidic oscillator having a source chamber communicating with saidinlet passage and an interaction chamber communicating with said sourcechamber, said pair of output ducts connected to said interactionchamber.

19. The device of claim 18 wherein each of said output ducts isconnected to at least one branch ending on the exterior surface of saidinner massive member and inside at least one of said alternate pistonchambers, and whereby at least one leakage path is provided from each.said piston chamber to a lower pressure region.

20. The device of claim 19 wherein said leakage paths converge and leadto said bit through said inner massive member.

21, The device of claim 19 wherein said leakage paths are orifices onthe outer wall of said outer massive member, discharging into the well,

22. The device of claim 18 wherein each said fluidic oscillator has apair of feedback paths, each of said feedback paths connecting saidinteraction chamber to a point of one of said output ducts to saidinteraction chamber.

23. The device of claim 18 wherein each said fluidic oscillator has apair of feedback paths, each of said feedback paths connecting saidinteraction chamber to a point on the exterior surfate of said innermassive member, said point lying within the gap between said exteriorsurface and said coacting flange.

24. The device of claim 18 wherein said output ducts of each saidfluidic oscillator are displaced relative to one another in thedirection of said longitudinal axis, the fluid flow through saidinteraction chamber and said output ducts being substantially transverseto said longitudinal axis, whereby axial acceleration of said innermassive member may act favorably toward switching the flow insaidinteraction chamber from one slid output duct to the other, resulting insaid resonant system sustaining itself in vibration through motionalfeedback 25. The device of claim 1 wherein said inlet passage connectsto an axial tubular cavity within said inner massive member internal tothe region of said piston chambers, the wall of said cavity beingtraversed by at least one pair of passages for each piston, the passagesof each pair being connectc l one'to-one to the piston chambers onopposite sides of said IlStOfl, and said oscillatory converting meansincludes a mechanical oscillator, said mechanical oscillator supportedwithin said tubular cavity by at least a pair of support members thatrave fluid sealing means and allow movement in one degrt of freedom,whereby said mechanical oscillator is allOWCll limited oscillatorymovement, said sealing means precluding direct access of said inlet flowof pressurized fluid to the inner openings of said passage pairs, saidmechanical oscillator being hollow and open to said fluid inlet passageat one end, the other end being blocked against flow, the wall of saidmechanical oscillator being travensed by holes forming a plurality ofnozzles, said nozzles being; located cooperatively adjacent to saidpassage pair inlets, whereby outward-extending jets may be formed bysaid pressurized fluid flowing through said nozzles and impinge upon oneor the other inlet of each said passage pair, causing pressure forces tobe developed alternately on one side or the other of each said piston,depending upon the relative instantaneous position of said mechanicaloscillator with respect to said inner massive member, such portingaction causing fluid energy to be imparted to said massive members atabout the natural frequency of said resonant system, the spent fluidfrom said piston cham bers and leakage flowing in the annular spacewithin said tubular cavity external to said mechanical oscillator andout through exit porting means.

26. The device of claim 25 wherein said support members are a pair oflaminate elastomer-metal bearings that have annular shape surroundingsaid mechanical oscillator, and have laminae parallel to saidlongitudinal axis, whereby limited axial movement of said oscillator isallowed.

27. The device of claim 25 wherein said support members have anassociated axial spring constant small enough in mag nitude to make thenatural frequency of axial resonance, defln :d by the mass of saidmechanical oscillator and said spring constant, low by comparison withthe natural frequency of said resonant system, whereby said mechanicaloscillator tends to remain virtually stationary as said inner massivemember reciprocates, causing said porting action to occur, and resultingin said resonant system sustaining itself in a vibratory state throughmotional feedback.

28. The device of claim 27 wherein additional damping means are providedat one end of said mechanical oscillator by the action ofa functionalpiston pumping fluid in and out of an enclosed volume through arestricted passage.

29. The device of claim 25 wherein said degree of freedom is along saidlongitudinal axis, further characterized by the presence of a smallfluid-operated oscillator within said inner massive member (i.e.,anykind of oscillator such as spool-type, diaphragm-type, fluidic type,etc.), said small fluid oscillator having a source chamber communicatingwith said internal inlet passage, and a small fluid-operated actuatorhaving at least one piston connected to said mechanical oscillator, atleast one output duct of said small fluid oscillator being connectedfunctionally to said small actuator to produce an oscillating axialforce on said pistons and to urge axial oscillation of same at about thefrequency of said resonant system resulting in said porting action bysaid mechanical oscillator.

30. The device of claim 29 wherein said small fluidoperated oscillatorhas an interaction chamber communicat' ing with said source chamber, anda pair of output ducts connected to said interaction chamber, saidoutput ducts being displaced relative to one another in the direction ofsaid longitudinal axis, the fluid flow through said interaction chamberand said output ducts being substantially transverse to saidlongitudinal axis, whereby axial acceleration of said inner massivemember may act favorably toward switching the flow in said interactionchamber from one said output duct to the other, said flow operating uponsaid small actuator resulting in said resonant system sustaining itselfin a vibratory state through motional feedback.

' 31. The device of claim 29 wherein the average magnitude of saidpiston forces becomes smaller than the average magnitude of inertialforces on said mechanical oscillator when the amplitude of oscillationof said inner massive member reaches a certain level, whereby control ofoscillation of said resonant system shifts to self-sustainment throughmotional feedback.

32. The device of claim 25 wherein said degree of freedom is around thelongitudinal axis, further characterized by the presence of a smallfluid-operated oscillator within said inner massive member, said smallfluid-operated oscillator having a source chamber communicating withsaid internal inlet passage, and a small angular fluid-operated actuatorfunctionally connected to said mechanical oscillator, at least oneoutput duct of said small fluid-operated oscillator being connectedfunctionally to said small actuator to produce an oscillating angularforce on said mechanical oscillator and to urge angular oscillation ofsame at about the frequency of said resonant system, resulting in saidporting action by said mechanical oscillator.

33. The device of claim 32 wherein said small fluidoperated oscillatoris of the fluidic type, and has an interaction chamber communicatingwith said source chamber, and a pair of output ducts connected to saidinteraction chamber, said output ducts being displaced axially relativeto one another, the fluid flow through said interaction chamber and saidoutput ducts being substantially transverse to said longitudinal axis,whereby axial acceleration of said inner massive member may actfavorably toward switching said flow in said interaction chamber fromone said output duct to the other, said flow operating upon said smallactuator, resulting in said resonant system sustaining itself throughmotional feedback.

34 The device of claim 32 wherein said support members are a pair oflaminated elastomer-metal bearings that have annular shape surroundingsaid mechanical oscillator, and have laminae inclined at an angle tosaid longitudinal axis, whereby limited relative angular movement ofsaid mechanical oscillator is allowed.

35. The device of claim 25 wherein said degree of freedom is along saidlongitudinal axis, further characterized by the presence of a fluiddamper including a piston and a cylinder member, one of said dampermembers attached to said mechanical oscillator and the other to one ofsaid massive members, whereby a damping force on said mechanicaloscillator results as a consequence of relative velocity between it andsaid massive member; and at least one axially operative spring attachedeffectively between said mechanical oscillator and the other of saidmassive members, whereby spring force results on said mechanicaloscillator as a consequence of relative displacement between it and saidother massive member, whereby the phase relationship of said mechanicaloscillator to said inner massive member depends upon said damping andspring forces.

36. The device of claim 35 wherein said support members are attached toand within said inner massive member and at least one of said supportmembers has axial spring properties whereby said axially-operativespring is effectively provided.

37. The device of claim 35 wherein one of said support members isattached within said outer massive member and has axial springproperties much stiffer than those of the other said support member,whereby said axiallyoperative spring is effectively provided.

38. The device of claim 1 and in addition a hydraulic turbine wheelconsisting of a plurality of radially positioned turbine blades rigidlyattached onto the interior wall surface of said outer massive member, aplurality of feeder duets provided in the body of said inner massivemember, said feeder ducts being connected to said internal inlet passageand directing high velocity jets of said pressurized fluid onto saidturbine blades, said high velocity jets transferring a substantial partof their momentum to said turbine wheel and urging said outer massivemember to rotate, conventional plain bearings inserted between saidouter massive member and said bearing members separating said outer andinner massive members, said conventional bearings permitting saidrotation of said outer massive member.

39. The device of claim 18 and in addition a hydraulic turbine wheelconsisting of a plurality of radially positioned tur bine blades rigidlyattached onto the interior wall surface of said outer massive member, aplurality of feeder ducts provided within the body of said inner massivemember, said feeder ducts being connected to one duct of said pair ofoutput ducts of said fluidic oscillator and directing high velocity jetsof said pressurized fluid onto said turbine blades, said high velocityjets transferring a substantial part of their momentum to said turbinewheel and urging said outer massive member to rotate, conventionalbearings inserted between said outer massive member and said bearingmembers separating said outer and inner massive members, saidconventional bearings permitting said rotation of said outer massivemember.

40. The device of claim 1 wherein at least one of said bearing membersis of the laminate elastomer-metal type consisting of alternate metaland elastomer laminae, said laminae having the shape of parallel helicalsurfaces, said shape resulting in said bearing members deformationbycompression having an axial component parallel to said longitudinal axisand an angular component about same said axis.

41. A vibratory drill assembly as described including:

an inner massive member at least partially telescoped within an outermassive member, the two massive members being separated by and heldconcentric to each other about a common longitudinal axis by two axiallyseparated laminate elastomer-metal bearing members as described, eachsaid bearing member having load faces inclined to said longitudinal axisand said bearing members being preloaded against one another to providestiff axial spring characteristics whereby limited axial displacement ofthe one massive member with respect to the other results upon theapplication of axial force between said massive members;

a bit drivingly attached to one of said massive members, and

a stem adapter attached resiliently in said axial direction to one ofsaid massive members; and

